As is known in the art and with reference to the enclosed FIG. 9 illustrates a compressor attached to the current state of the art, reciprocating piston compressors generate pressure by compressing a gas within a cylinder 30 through the axial movement of a piston 29, and the gas of the low pressure side (suction or evaporation pressure) enters in the interior of the cylinder through a suction valve 32, is compressed within the cylinder 30 by movement of piston 29 and then leaves the cylinder by a discharge valve 33, moving to the high pressure side (discharge or condensation pressure).
In the case of resonant linear compressors, the piston is driven by a linear actuator formed by a support 34 and magnets 35 (which can be actuated by one or more coils 36), and one or more springs 38, 39 connect the movable part (piston, supports and magnets) to the fixed part (cylinder, stator 44, coil, head 31 and structure 45). The movable parts and the springs form the resonant assembly of the compressor.
Then, the resonant assembly driven by linear motor has the function to develop a linear reciprocating motion, causing the movement of the piston inside the cylinder to exert a compression action of the gas admitted by the suction valve, until the point at which it can be discharged through the discharge valve to the high pressure side.
The amplitude of the operation of the linear compressor is regulated by the balance between the power generated by the motor and the power consumed by the compression mechanism, plus the losses generated in this process. To achieve maximum thermodynamic efficiency and maximum cooling capacity, it is necessary that the maximum displacement of the piston approaches the maximum possible to the end of stroke (head), thus reducing the dead volume of gas in the compression process.
To enable this process, it is necessary that the stroke of the piston be known with great accuracy to avoid the risk of impact of the piston with the end of the stroke (head 31), since this impact can generate from acoustic noise and loss of efficiency to compressor failure. Thus, the higher the error in the estimation measurement of the position of the piston, the greater will be the safety coefficient necessary between the maximum displacement and the end of the stroke, so that it is possible to safely operate the compressor—leading to loss of performance of the product.
However, if it is necessary to reduce the cooling capacity of the compressor due to less need of the refrigeration system, it is possible to reduce the maximum stroke of the piston operation which consequently reduces the power supplied to the compressor thus, it is possible to control the capability of the refrigeration compressor, obtaining a variable capacity.
Another important feature of the resonant linear compressors is the driving frequency. These systems are designed to operate at resonant frequency of the system mass/spring, a condition in which efficiency is maximal, where the mass (m) is the sum of the mass of the components of the movable part (piston, support and magnets), and the equivalent spring (KT) is the sum of the resonant spring of the system (KML) with the gas spring generated by the compression force of the gas (KG) which has a behavior similar to a variable and nonlinear spring, which depends on the evaporation and condensation pressure of the cooling system and also the gas used in the system. The resonance frequency (fR) can be calculated from equations (1) and (2) given below:
                              K          T                =                              K            ML                    +                      K            G                                              (        1        )                                          f          R                =                              1                          2              ·              π                                ⁢                                                    K                T                            m                                                          (        2        )            
Due to the portion of gas spring (KG)—which is unknown, nonlinear and variable throughout the operation—it is not possible to calculate the resonance frequency with the precision necessary to optimize the efficiency of the compressor. In another alternative way of adjusting the resonant frequency, it is applied a variation in the driving frequency until it is in the point of maximum power for a constant current. This method is simple and easy to implement, but its drawback is the fact that it is necessary to disrupt the system periodically to detect the resonant frequency.
When the system operates in resonant frequency, the motor current is in quadrature with the displacement or the motor current is in phase with the back electromotive force of the motor (FCEM), as FCEM is proportional to the derivative of displacement. This method is more accurate, but it requires the measurement of the current phase and the displacement phase or of FCEM, having the disadvantage of requiring the installation of sensors of position or speed.
An alternative construction to the resonating linear compressor is proposed in the patent application PI 0601645-6 which aims to reduce vibration of the compressor and also the size and weight, wherein in this construction the piston is connected to the actuator by the resonant spring generating two movable parts in relation to the structure of the compressor, thereby increasing the difficulty of controlling the mechanical engine due to the need to monitor and control two movable parts. However, also for this compressor with two movable parts it is necessary to monitor the phase of the actuator speed in relation to the current phase and the maximum stroke of the piston.
Other proposed solutions to obtain the stroke of the compressor involve the use of position sensors, such as those described in the following documents:                PI 0001404-4 (EMBRACO)—describes a sensor impeller, which has the disadvantages of the difficulty of insulation and noise of the electrical contact;        PI 0203724-6 (EMBRACO)—describes an inductive sensor mounted on the valves board, which allows to measure the distance piston plate directly on top of the piston. It is a solution of high precision, but requires a space for installing the sensor in the valves board and, moreover, has a higher cost and accurate calibration;        PI 0301969-1 (EMBRACO)—provides the use of PZT sensor, operating similar to an accelerometer; which has a good sensitivity for detection of impact, but it has a greater error in the measurement of the position;        PI PI 0704947-1 and 0705049-6 (both of EMBRACO)—provides a coil installed inside the engine to monitor the movement of the magnet of the linear actuator; it needs a time without current in the engine, in the region of measurement of the maximum stroke, which thereby limits the maximum power and flexibility of control of the equipment.        U.S. Pat. No. 5,897,296 and JP 1,336,661 (Sanyo)—use sensor, A/D converter and discrete/digital signal, and subsequently interpolated to determine the maximum forward position of the piston. With this solution, it is possible to achieve a high degree of accuracy, but the measurement is not made at the site of interest (distance piston/plate), whereby there is the need of considering the tolerances of the mounting position of the transducer and possible need of calibration. It also has the disadvantage of presenting a high cost;        U.S. Pat. No. 5,897,269 (Matsushita)—performs the control with position sensor, it presents possible need for calibration and high cost.        
All the above solutions have been developed for a system with a movable part and using the position sensor and thus are not suitable for the compressor with two movable parts.
Other solutions that do not use the position sensor are described in documents:                U.S. Pat. No. 5,342,176, U.S. Pat. No. 5,496,153, U.S. Pat. No. 4,642,547 (Sunpower) and U.S. Pat. No. 6,176,683, KR 96-79125 and KR 96-15062 (LG)—They carry out the calculation of speed from the electric equation, and with speed the stroke is calculated; this method is not accurate because it considers the dynamics of the compressor and does not estimate the offset of the stroke;        WO 00079671 (F & P)—the operating limit is calculated from a table between the resonance frequency and the evaporation temperature; as a disadvantage, this method does not have good precision and requires a temperature or pressure sensor.        WO 03044365 (F & P)—the limit of operation is obtained by detecting the impact, by varying the resonance frequency of the compressor, this method has the disadvantage of generating acoustic noise and oscillation of the stroke at maximum capacity.        
The above solutions without position sensor, were also designed for a system with a movable part, and thus are not suitable for the compressor with two movable parts.
Solutions to the problem of the drive frequency are suggested in the documents indicated below:                WO 00079671A1 (F & P)—uses detecting back electromotive force of the motor to adjust the resonance frequency. This technique has the disadvantage of needing a minimum time without current in order to detect the zero crossing of FCEM; thus affecting the maximum power and the efficiency by the distortion in the waveform of the current.        U.S. Pat. No. 5,897,296 (Matsushita)—Control with position sensor and control of frequency to minimize the current. This technique has the disadvantage of requiring to disturb the system periodically to adjust the driving frequency—which can impair product performance.        U.S. Pat. No. 6,832,898 (Matsushita)—Control of the operation frequency by the maximum power for a constant current. This technique uses the same principle of the previous citation, so has the same disadvantage of requiring to disturb the system periodically.        U.S. Pat. No. 5,980,211 (Sanyo)—Control with sensor and control of frequency by the phase with the position. This method has the disadvantage of needing a position sensor.        
In short, in the current state of the art it is necessary to use two position sensors to control the compressor with two movable parts, and control techniques without sensors were not developed for this type of compressor.